High performance heat exchange assembly

ABSTRACT

Heat sinks are provided that achieve very high convective heat transfer surface per unit volume. These heat sinks comprise a spreader plate, at least two fins and at least one porous reticulated foam block that fills the space between the fins.

CROSS-REFERENCE TO RELATED APPLICATIONS

[0001] This application claims the benefit of U.S. Provisional PatentApplication Ser. No. 60/189,133, filed Mar. 14, 2000.

TECHNICAL FIELD

[0002] The present invention is directed to heat sinks primarily for usein dissipating waste heat generated by electrical and/or electroniccomponents and assemblies. These heat sinks include a heat spreaderplate and an assembly of heat conducting fins and reticulated foamstructures that are bonded together. Electronic components are connectedto one surface of the spreader plate with the assembly of fins and foamconnected to another surface of the spreader plate in contact with acooling fluid.

BACKGROUND OF THE INVENTION

[0003] High power electrical and electronic components continue to havean increasing demand for higher power dissipation within a relativelyconfined space. In order to provide for such higher power dissipationrequirements while remaining suitably compact, several levels of thermalmanagement are usually required at the device, sub-assembly andcomponent level.

[0004] At the component level, various types of heat exchangers and heatsinks have been used that apply natural or forced convection or othercooling methods. A typical heat sink for electrical or electroniccomponents is depicted in FIG. 1. As shown, the heat sink includes aheat spreader plate 10 to which metal fins 12 are attached. Anelectronic component is attached to face 14 of spreader plate 10 and acooling fluid 16, such as air or water, is passed across fins 12 todissipate the heat generated by the electronic component. For a givenpower level to be dissipated, the spreader plate size (i.e., area) andthe fin length along the length of the cooling flow path can becalculated using known methods. Fin spacing and fin height are usuallydetermined by known methods such as numerical modeling.

[0005] In demanding applications, the method of heat exchange is usuallyforced convection to the cooling fluid. In such systems, heat exchangecan be improved by increasing the fin surface area exposed to thecooling fluid. This is accomplished by increasing the number of the finsper unit volume. However, there are limitations to achievable findensities based upon manufacturing constraints and cooling fluid flowrequirements.

[0006] Reticulated foams are also known in the art for their ability toconduct heat such as the metal foams disclosed in U.S. Pat. Nos.3,616,841 and 3,946,039 to Walz, and the ceramic foams disclosed in U.S.Pat. No. 4,808,558 to Park et al. Metal foams have been sold under thetrade name DUOCEL available from Energy Research and Generation, Inc.,Oakland, Calif.

[0007] Until recently, metal and ceramic reticulated foams have not beenadapted for use in heat sinks for dissipating waste heat from electroniccomponents. However, these structures, especially when comprised ofmetal, make excellent heat exchangers because of their conductivity andtheir extremely high surface area to volume ratio. While earlier porousheat exchangers had up to 100 open cells per square inch, reticulatedfoam has up to 15,625 open cells per square inch. Reticulated foam isfar more porous and has far more surface area per unit volume (1600square feet/cubic foot) than heat exchangers having other structures.The pressure drop of fluids flowing through reticulated foam is alsorelatively low so that movement of a cooling fluid through the foam ispractical.

[0008] Studies by Bastawros have now shown the efficacy of metallicfoams in forced convection heat removal for cooling of electronics. See,Bastawros, A. -F., 1998, Effectiveness of Open-Cell Metallic Foams forHigh Power Electronic Cooling, ASME Conf. Proc. HTD-361-3/PID-3,211-217, and Bastawros, A. -F., Evans, A. G. and Stone, H. A., 1998,Evaluation of Cellular Metal Heat Transfer Media, Harvard Universityreport MECH 325, Cambridge, Mass. Bastawros demonstrated that the use ofmetallic foam improved heat removal rate with a moderate increase in thepressure drop. Bastawros' results were based on thermal and hydraulicmeasurements (on an open cell aluminum alloy foam having a pore size of30 pores per inch) used in conjunction with a model based upon a bank ofcylinders in cross-flow to understand the effect of various foammorphologies. The model prediction was extrapolated to examine thetrade-off between heat removal and pressure drop. The measurementsshowed that a high performance cellular aluminum heat sink (i.e.,aluminum foam) removed 2-3 times the usual heat flux removed by apin-fin array with only a moderate increase in pressure drop.

SUMMARY OF THE INVENTION

[0009] A range of new heat sinks for electrical and electroniccomponents is herein presented that provides for space-efficient heatexchange with low thermal resistance. These heat sinks are capable ofremoving the increased waste heat flux generated by today's higher powerelectronic systems.

[0010] In general, heat sinks of the present invention comprise aspreader plate, at least two fins and at least one porous reticulatedfoam block that fills the space between the fins. All materials are madefrom a heat conducting material. The fins and foam blocks form anassembly that is connected to one surface of the spreader plate.Electronic components to be cooled are preferably connected to anopposing surface of the spreader plate, but may be connected to anysurface of the spreader plate suited for heat transfer.

[0011] The present invention further defines the preferred dimensionalrelationships for establishing the optimum fin spacing and fin heightfor the heat sinks provided herein. Devices produced using thesedimensional relationships find particular use in cooling microelectroniccomponents such as microprocessors.

BRIEF DESCRIPTION OF THE DRAWINGS

[0012]FIG. 1 shows a typical heat sink of the prior art.

[0013]FIGS. 2a and 2 b show the front and top views of the improved heatsink of the present invention.

DESCRIPTION OF THE PREFERRED EMBODIMENT

[0014] By the present invention, it has been discovered that heat sinksthat use a combination of solid non-porous fins and highly porousreticulated foam can provide improved performance over knownapplications that use one or the other. It is fully contemplated thatany combinations of fins and reticulated foam may be used in a widevariety of different applications to achieve improved cooling.

[0015] It has been further discovered that there are constraints on thevolume and geometry of reticulated foam beyond which the use ofadditional foam will have little impact on the overall heat sink'sability to dissipate thermal power at a given flow rate (i.e., theperformance). This is because the reticulated foam is not a fully densematerial (e.g., aluminum foam has a density of about 10% of solidaluminum). Therefore, a high convective heat transfer surface area isachieved at the expense of reduced thermal conductivity.

[0016] Additionally, in microelectronic cooling applications such as forexample for microprocessors, practical considerations relative topackage size, air flow rate, pressure drop and noise limits can imposefurther constraints on possible configurations and dimensions.Nonetheless, using the methods of the present invention, suitable heatsinks can be produced.

[0017] Heat sinks of the present invention achieve very high convectiveheat transfer surface per unit volume. These heat sinks comprise aspreader plate, at least two fins and at least one porous reticulatedfoam block that fill the space between the fins. This basic structuremay be expanded to any configuration comprising foam blocks in betweenat least two fins that are mounted onto surface of a spreader plate.

[0018] Primary heat transfer to the cooling fluid is by convection fromthe foam, with the fins and spreader plate being used to conduct heatfrom the connected heat source (i.e., the electronic component) into thefoam. In air-to-air heat exchange (i.e., where air is being used as thecooling fluid), ambient air may be drawn in through the foam's openvertical side walls and exhausted through the foam's top surface, orvice versa.

[0019] A preferred embodiment of the present invention is shown in FIG.2a and FIG. 2b. As shown, the device comprises a heat spreader plate 20,with fins 22 and reticulated foam blocks 24 filling the space in-betweenthe fins. The fins 22 and foam blocks 24 form an assembly that ismounted onto one surface of the spreader plate 20, leaving an opposingsurface free for contact with an electronic component to be cooled.

[0020] Referring to FIG. 2a, fins 22 are mounted so that they aresubstantially perpendicular to the spreader plate 20. Foam blocks 24 aremounted in-between fins 22 to fill the width region that defines thehorizontal space between adjacent fins. The foam blocks 24 alsopreferably fill the height region that defines the vertical spacebetween adjacent fins to the height of the fins 22. While FIG. 2a showsthat the foam blocks 24 fill the height region, it is contemplated thatin alternative embodiments the foam blocks may partially fill oroverfill the height region.

[0021] Referring to FIG. 2b, heat sinks of the present invention areconfigured such that the fins 22 are substantially parallel to oneanother and aligned along the length of the spreader plate 20.

[0022] The selection of spreader plate size and fin length along thecooling flow length, for a given power dissipation requirement, followthose techniques known in the art. The overall dimensions of thespreader plate are generally fixed based on the amount of heat to bedissipated from the surface of the heat source (such as a computerchip). The spreader plate surface area should be such that, for aprescribed flow rate of the cooling fluid flowing over the spreaderplate, the heat from the heat source is able to spread to the edges ofthe spreader plate. Additional considerations may also be determinativeof the spreader plate surface area such as packaging constraints.

[0023] Generally, however, spreader plate surface area is selected bymultiplying the surface area of the heat source with the areamagnification factor. The area magnification factor λ=A_(s)/A_(h)represents a ratio of the surface area of the heat source A_(h) with thesurface area of the spreader plate A_(s). Typical values of λ are in therange of 8 to 12, and are generally used in calculating spreader platesurface area for a given surface area of a heat source. From thestandpoint of heat removal efficiency, λ should be as low as possible.Highly effective heat transfer surfaces such as highly conductive finsof optimized dimensions and/or the use of heat transfer augmentationmeans such as reticulated foam provide for relatively low values of λ.For example, in the present invention, if the surface area of the heatsource is 1.5 in² and the selected area magnification factor is taken asλ=8 (for highly efficient transfer), then the surface area of thespreader plate will be 8×1.5=12 in². For a spreader plate of this area,packaging considerations could prescribe a length of the plate in theflow direction to be 4 in. Then the width of the spreader plate will be3 in.

[0024] The number of fins and the fin length in the flow direction canbe determined following calculation of fin spacing. For example, basedon the applicable heat transfer considerations, if the fin spacinga=0.5938 in. and the fin thickness δ_(f)=0.125 in., then the maximumnumber of fins n can be determined as follows. The width of the spreaderplate is equal to the space occupied by n fins, nδ_(f), plus the spacetaken up by the inter-fin gaps, (n−1)a. For a spreader plate having awidth of 3 inches, the number of fins can be calculated from theequation, nδ_(f)+(n−1)a=3 in. Using δ_(f)=0.125 in., a=0.5938 in., andsolving for n, the number of fins is equal to 5.

[0025] In preferred heat sinks of the present invention, the fin spacingand fin (and foam) height are optimized according to the followingformulas. Based on heat transfer considerations, the optimum fin height(as shown in FIG. 2a), b, is determined using the relation,$\begin{matrix}{b = {0.6498\sqrt{\frac{k_{f}\delta_{f}}{h}}}} & (1)\end{matrix}$

[0026] where,

[0027] k_(f) is the thermal conductivity of the selected fin material,Btu/ft s ° F.

[0028] δ_(f) is the fin thickness, ft

[0029] h is the convective heat transfer coefficient for the foam-filledspace bounded by the fins and the spreader plate, Btu/ft² s ° F., andwhere h is given by the formula, $\begin{matrix}{h = {{1.2704\left\lbrack \frac{n^{0.50}}{\left( {1 - \varphi} \right)^{0\quad 25}} \right\rbrack}\left( \frac{\rho^{0.50}k^{0.63}c_{p}^{0.37}}{\mu^{0.13}} \right)u_{m}^{0\quad 50}}} & (2)\end{matrix}$

[0030] where,

[0031] n is the linear density of the foam material, pores per ft

[0032] φ is the foam porosity expressed as a fraction

[0033] ρ is the density of the flowing fluid, lb_(m)/ft³

[0034] k is the thermal conductivity of the flowing fluid, Btu/ft s ° F.

[0035] c_(p) is the isobaric specific heat of the flowing fluid,Btu/lb_(m) ° F. μ is the dynamic viscosity of the flowing fluid,lb_(m)/ft s

[0036] u_(m) is the mean velocity of the flowing fluid, ft/s

[0037] According to the present invention, the optimum fin spacing (asshown in FIG. 2a), a, is determined by the relationship,

a=Φδ  (3)

[0038] where,

[0039] Φ is a numerical factor whose values, determined by numericalexperimentation and measurements, range between about 1 to about 6.

[0040] δ is the minimum fin spacing, ft

[0041] The preferred value of Φ is about 2.5. However, depending on heatdissipation and pressure drop considerations, higher values of Φ may beemployed.

[0042] Based on heat transfer considerations, the minimum fin spacing δis determined by the relation, $\begin{matrix}{\delta = {7.32\sqrt{\frac{kc}{\rho \quad c_{p}u_{m}}}}} & (4)\end{matrix}$

[0043] where,

[0044] c is the fin length in the flow direction (as shown in FIG. 26),ft

[0045] k is the thermal conductivity of the flowing fluid, Btu/ft s ° F.

[0046] ρ is the density of the flowing fluid lb_(m)/ft³

[0047] c_(p) is the isobaric specific heat of the flowing fluid,Btu/lb_(m) ° F.

[0048] u_(m) is the mean velocity of the flowing fluid, ft/s.

[0049] To maximize thermal conduction from the heat source through thespreader plate into the fins and foam, the fins and foam are bonded toone another and the spreader plate. While thermal bonding such asbrazing is preferred, any suitable means may be employed including, forexample, using a conductive epoxy to form an adhesive bond. In preferredheat sinks of the present invention, the fins, the foam blocks and thespreader plate are assembled and then preferably furnace-brazed to bondthe foam to the fins and the spreader plate.

[0050] The spreader plate and fins are solid and made from thermalconducting materials known in the art. The reticulated foam is an opencell media and also made from known thermal conducting materials.Preferred thermal conducting materials include aluminum, copper,graphite and aluminum-nitride ceramics. The spreader plate, fins and thereticulated foam may be selected from the same or different materials.In a preferred embodiment, the selected thermal conducting material forthe spreader plate, fins and the reticulated foam is aluminum.

[0051] The reticulated foam structure is available from commercialsources or may be made by methods known in the art. Suitable processesfor making metal foams are disclosed in U.S. Pat. Nos. 3,616,841 and3,946,039 to Walz, and processes for making ceramic foams are disclosedin U.S. Pat. No. 4,808,558, the teachings of which are incorporatedherein by reference. Reticulated foam metal can be formed by themanufacturer to have many shapes, densities and cell sizes. Foam blocksas used herein may be obtained from such manufacturers or cut fromlarger pieces. Aluminum foams suitable for use herein are availableunder the tradename DUOCEL from Energy Research and Generation, Inc.,Oakland, Calif.

[0052] The following examples are provided to illustrate heat sinks ofthe present invention designed for microelectronic cooling applicationsusing the relationships set forth above and based upon a powerdissipation requirement of up to about 200 watts.

EXAMPLE 1

[0053] In a heat sink of the present invention, aluminum fins areselected having a thickness δ_(f)=0.125 inch (0.0104 ft) with thermalconductivity k_(f)=133 Btu/ft hr° F. (0.0369 Btu/ft s ° F.). The finlength c in the flow direction, dictated by the packaging and heatdissipation consideration, is 4 inches (0.3333 ft). A commerciallyavailable open cell aluminum foam with linear density of n=20 pores perinch (240 pores/ft) and a porosity φ=0.90 is also selected. The coolingmedium is ambient air flowing with a mean velocity u_(m)=10 ft/s. Thetransport properties of the ambient air are as follows:

[0054] Density ρ=0.0749 lb_(m)/ft³

[0055] Thermal conductivity k=0.0000041 Btu/ft s ° F.

[0056] Isobaric specific heat c_(p)=0.2410 Btu/lb_(m)° F.

[0057] Dynamic viscosity μ=0.0000123 lb_(m)/ft s

[0058] To determine optimum fin height b, the convective heat transfercoefficient h is first determined using Equation (2), above, providingh=0.0313 Btu/ft²s ° F. Next, introducing this value of h into Equation(1), above, we obtain the optimum fin height b=0.0721 ft. (0.8646 inch).

[0059] To determine the optimum fin spacing a, the minimum fin spacing δis first determined using Equation (4), to obtain δ=0.0201 ft. (0.2417inch). Then, using Equation (3) and selecting the factor Φ=2.5 resultsin an optimum fin spacing a=0.6043 inch.

EXAMPLE 2

[0060] This example is the same as Example 1 except that the finmaterial has been changed from aluminum to copper. The copper fins havea thickness δ_(f)=0.125 inch (0.0104 ft) and thermal conductivityk_(f)=226 Btu/ft hr ° F. (0.0628 Btu/ft s ° F.). The fin length c in theflow direction, dictated by the packaging and heat dissipationconsideration, is 4 inches (0.3333 ft). The reticulated foam is acommercially available open cell aluminum foam having a linear densityn=20 pores per inch (240 pores/ft) and a porosity φ=0.90. The coolingmedium is ambient air flowing with a mean velocity u_(m)=10 ft/s. Thetransport properties of the ambient air are as follows.

[0061] Density ρ=0.0749 lb_(m)/ft³

[0062] Thermal conductivity k=0.0000041 Btu/ft s ° F.

[0063] Isobaric specific heat c_(p)=0.2410 Btu/lb_(m)° F.

[0064] Dynamic viscosity μ=0.0000123 lb_(m)/ft s

[0065] As in Example 1, using Equation (2), the convective heat transfercoefficient h=0.0313 Btu/ft² s ° F. Then, using Equation (1), we obtainthe optimum fin height b=0.0939 ft (1.1271 inch). This optimal height ofthe copper fin is 30% higher than that for the aluminum fin indicatingthat for the same fin thickness, the copper fin has higher heatdissipation potential than aluminum fin. This can be attributed to thehigher thermal conductivity of copper.

[0066] Optimum fin spacing a follows Example 1. The minimum fin spacingδ is first determined using Equation (4), to obtain δ=0.0201 ft. (0.2417inch). Then, using Equation (3) and selecting the factor Φ=2.5 resultsin an optimum fin spacing a=0.6043 inch.

[0067] While the preferred embodiment of the present invention has beendescribed so as to enable one skilled in the art to practice the heatsinks disclosed, it is to be understood that variations andmodifications may be employed without departing from the concept andintent of the present invention as defined by the following claims. Thepreceding description and examples are intended to by exemplary andshould not be read to limit the scope of the invention. The scope of theinvention should be determined only by reference to the followingclaims.

1. A heat sink for electrical or electronic components comprising: aheat spreader plate to which the components to be cooled are connected;at least two heat conducting fins that are positioned substantiallyparallel to one another and which are connected substantiallyperpendicular to said heat spreader plate; and highly porous heatconducting reticulated foam block that fills the space between parallelfins.
 2. A heat sink of claim 1 wherein said fins and said foam blocksare connected to one surface of said heat spreader plate.
 3. A heat sinkof claim 1 wherein the fin height, b, is determined by the relationship,${b = {0.6498\sqrt{\frac{k_{f}\delta_{f}}{h}}}}\quad$

where, k_(f) is the thermal conductivity of the selected fin material,Btu/ft s ° F. δ_(f) is the fin thickness, ft h is the convective heattransfer coefficient for the foam-filled space bounded by the fins andthe spreader plate, Btu/ft² s ° F., and where h is given by the formula,$h = {{1.2704\left\lbrack \frac{n^{0\quad 50}}{\left( {1 - \varphi} \right)^{0.25}} \right\rbrack}\left( \frac{\rho^{{0\quad 50}\quad}k^{0.63}c_{p}^{0\quad 37}}{\mu^{0.13}} \right)u_{m}^{0.50}}$

where, n is the linear density of the foam material, pores per ft φ isthe foam porosity expressed as a fraction ρ is the density of theflowing fluid, lb_(m)/ft³ k is the thermal conductivity of the flowingfluid, Btu/ft s ° F. c_(p) is the isobaric specific heat of the flowingfluid, Btu/lb_(m) ° F. μ is the dynamic viscosity of the flowing fluid,lb_(m)/ft s u_(m) is the mean velocity of the flowing fluid, ft/s
 4. Aheat sink of claim 1 wherein the fin spacing, a, is determined by therelationship, a=Φδ where, Φ is between 1 to 6 δ, ft, is determined bythe relation,$\delta = {7.32\sqrt{\frac{kc}{\rho \quad c_{p}u_{m}}}}$

where, c is the selected fin length in the flow direction, ft k is thethermal conductivity of the flowing fluid, Btu/ft s ° F. ρ is thedensity of the flowing fluid lb_(m)/ft³ c_(p) is the isobaric specificheat of the flowing fluid, Btu/lb_(m)° F. u_(m) is the mean velocity ofthe flowing fluid, ft/s.
 5. A heat sink of claim 1 wherein said heatspreader plate, said fins and said heat conducting foam are made fromthe same or different thermal conducting materials.
 6. A heat sink ofclaim 1 wherein said heat spreader plate, said fins and said heatconducting foam are made from aluminum, copper, graphite oraluminum-nitride ceramic.
 7. A heat sink of claim 1 wherein said heatspreader plate, said fins and said heat conducting foam are made fromaluminum.
 8. A method of making a heat sink comprising a heat spreaderplate, at least two fins and reticulated foam block that fills the spacein-between the fins comprising, selecting said heat spreader plate, saidfins and said foam block; assembling said fins and said foam block ontosaid spreader plate so that said fins are substantially parallel to oneanother and substantially perpendicular to said spreader plate and saidfoam block fill the space in between said fins; and bonding the assemblyof said fins and said foam block to said spreader plate.
 9. A method ofclaim 8 wherein the assembly of said fins and said foam block areconnected to one surface of said heat spreader plate.
 10. A method ofclaim 8 wherein said bonding is accomplished using a thermallyconductive adhesive or furnace brazing.
 11. A method of claim 8 whereinthe fin height, b, is determined by the relationship,$b = {0.6498\sqrt{\frac{k_{f}\delta_{f}}{h}}}$

where, k_(f) is the thermal conductivity of the selected fin material,Btu/ft s ° F. δ_(f) is the fin thickness, ft h is the convective heattransfer coefficient for the foam-filled space bounded by the fins andthe spreader plate, Btu/ft² s ° F., and where h is given by the formula,$h = {{1.2704\left\lbrack \frac{n^{0\quad 50}}{\left( {1 - \varphi} \right)^{0.25}} \right\rbrack}\left( \frac{\rho^{{0\quad 50}\quad}k^{0.63}c_{p}^{0\quad 37}}{\mu^{0.13}} \right)u_{m}^{0.50}}$

where, n is the linear density of the foam material, pores per ft φ isthe foam porosity expressed as a fraction ρ is the density of theflowing fluid, lb_(m)/ft³ k is the thermal conductivity of the flowingfluid, Btu/ft s ° F. c_(p) is the isobaric specific heat of the flowingfluid, Btu/lb_(m) ° F. μ is the dynamic viscosity of the flowing fluid,lb_(m)/ft s u_(m) is the mean velocity of the flowing fluid, ft/s
 12. Amethod of claim 8 wherein the fin spacing, a, is determined by therelationship, a=Φδ where, Φ is between 1 to 6 δ, ft, is determined bythe relation,$\delta = {7.32\sqrt{\frac{kc}{\rho \quad c_{p}u_{m}}}}$

where, c is the selected fin length in the flow direction, ft k is thethermal conductivity of the flowing fluid, Btu/ft s OF ρ is the densityof the flowing fluid lb_(m)/ft³ c_(p) is the isobaric specific heat ofthe flowing fluid, Btu/lb_(m)° F. u_(m) is the mean velocity of theflowing fluid, ft/s.
 13. A method of claim 8 wherein said heat spreaderplate, said fins and said heat conducting foam are made from the same ordifferent thermal conducting materials.
 14. A method of claim 8 whereinsaid heat spreader plate, said fins and said heat conducting foam aremade from aluminum, copper, graphite or aluminum-nitride ceramic.
 15. Amethod of claim 8 wherein said heat spreader plate, said fins and saidheat conducting foam are made from aluminum.
 16. A method of coolingelectronic components by attaching the electronic components to onesurface of a heat sink and passing a cooling fluid over the opposingsurface of the heat sink, wherein said heat sink comprises, a heatspreader plate, at least two heat conducting fins that are positionedsubstantially parallel to one another and which are connectedsubstantially perpendicular to said heat spreader plate, and highlyporous heat conducting reticulated foam block that fills the spacebetween parallel fins, wherein the height of said fins, b, and isdetermined by the relationship,$b = {0.6498\sqrt{\frac{k_{f}\delta_{f}}{h}}}$

where, k_(f) is the thermal conductivity of the selected fin material,Btu/ft s ° F. δ_(f) is the fin thickness, ft h is the convective heattransfer coefficient for the foam-filled space bounded by the fins andthe spreader plate, Btu/ft² s ° F., and where h is given by the formula,$h = {{1.2704\quad\left\lbrack \frac{n^{0.50}}{\left( {1 - \varphi} \right)^{0.25}} \right\rbrack}\quad \left( \frac{\rho^{0.50}k^{0.63}c_{p}^{0.37}}{\mu^{0.13}} \right)u_{m}^{0.50}}$

where, n is the linear density of the foam material, pores per ft Φ isthe foam porosity expressed as a fraction ρ is the density of theflowing fluid, lb_(m)ft³ k is the thermal conductivity of the flowingfluid, Btu/ft s ° F. c_(p) is the isobaric specific heat of the flowingfluid, Btu/lb_(m) ° F. μ is the dynamic viscosity of the flowing fluid,lb_(m)/ft s u_(m) is the mean velocity of the flowing fluid, ft/s andwherein the fin spacing, a, is determined by the relationship, a=Φδwhere, Φ is between 1 to 6 6, ft, is determined by the relation,$\delta = {7.32\quad \sqrt{\frac{kc}{\rho \quad c_{p}u_{m}}}}$

where, c is the selected fin length in the flow direction, ft k is thethermal conductivity of the flowing fluid, Btu/ft s ° F. ρ is thedensity of the flowing fluid lb_(m)ft³ c_(p) is the isobaric specificheat of the flowing fluid, Btu/lb_(m)° F. u_(m) is the mean velocity ofthe flowing fluid, ft/s.
 17. A method of claim 16 wherein the electroniccomponent is a microprocessor, the cooling fluid is air and the heatsink is made from aluminum materials.
 18. A method of claim 17 whereinthe air is drawn in from the open side walls of said foam blocks andexhausted out of the top of said foam blocks.
 19. A method of claim 17wherein the air is drawn through said foam blocks along the entirelength of said parallel fins.
 20. A method of claim 17 wherein the airis pushed through said foam blocks.